Torsional vibration dampers

ABSTRACT

A torsional vibration damper such as a twin mass flywheel ( 10 ), having an input clement ( 11 ) and an output element ( 12 ) which are relatively rotatable against the action of a damping means which includes a plurality of linkages( 40 ). Each linkage has one or more links ( 41 ) mounted on one element ( 11 ) via pivot pins ( 43 ) which are arranged to extend generally radially in use and connected with the other element ( 12 ) by one or more additional links( 42 ). The pivot pins ( 43 ) of the generally radially extending links ( 41 ) are supported from the associated damper element ( 11 ) via spaced bearings (43 a , 43 b ). A number of Such spaced bearing arrangements are disclosed some of which hold grease captive.

FIELD OF THE INVENTION

This invention relates to torsional vibration dampers such as twin massflywheels and in particular such flywheels used to connect an internalcombustion engine with a gearbox on a motor vehicle.

BACKGROUND OF THE INVENTION

In, for example, the Applicant's earlier granted patents GB 2229793 and2282868 and pending applications PCT/GB95/01976 and PCT/GB97/00361 thereare described torsional vibration dampers in the form of twin massflywheels in which an input element and an output element are relativelyrotatable against the action of a damping means which includes aplurality of linkages each linkage including one or more link memberswhich act as bob weights to provide a speed dependant damping on therelative rotation of the flywheel elements. The Applicant's earlierpatent applications GB-A-2220464 and PCT/GB97/30302 also disclosetorsional vibration dampers used in combination with a fluid coupling inwhich the relative rotation of input and output elements of the damperis again controlled by a plurality of such linkages.

Such vibration dampers can be prone to problems associated with tiltingof the input element out of its normal plane of rotation due to flexingof the crankshaft of the associated engine. Such tilting can imposeundesirable loadings in the linkages which interconnect the elements.

It is an object of the present invention to provide a vibration damperwhich at least partially mitigates the above problem.

SUMMARY OF THE INVENTION

Thus according to the present invention there is provided a torsionalvibration damper comprising an input element and an output element whichare relatively rotatable against the action of a damping means whichincludes a plurality of linkages, each linkage comprising one or morelinks mounted on one element via pivot pins and arranged to extendgenerally radially in use and connected with the other element by one ormore additional links, damper being characterised in that the pivot pinsof the generally radially extending links are supported from said onedamper element via spaced bearings in said one damper element.

Such a spaced bearing arrangement is applicable to all the torsionalvibration dampers disclosed in the above referred to earlier patents andapplications when used as a twin mass flywheel where the input andoutput elements comprise input and output flywheel masses respectivelyor when used as a damper in which the input and output elements are ofrelatively light mass and are connected in a drive-line either as aseparate unit or in combination with a fluid coupling as discussedabove.

Such a torsional vibration damper may also include generally radiallyextending links which have an intermediate portion designed to flex inan axial sense to accommodate forces applied to the link via theassociated additional link as a result of relative tilting of the damperelements.

The pins which mount the generally radially extending links arepreferably a clearance fit in the bearings, the level of clearance beingsufficient to ensure that any thermal expansion of the pins during useof the flywheel is insufficient to cause the pins to bind in thebearings.

For example, when a mounting pin of 10 mm diameter is used a clearanceof 0.1 m (100 microns) between the internal diameter of the bearings andthe external diameter of the mounting pin is sufficient (i.e. aclearance of the order of 1% of pin diameter).

The bearings can be in the form of plain polymeric bushes (for examplepolyimide resin with a filler such as graphite). Many other alternativeplain bearing bush materials could be used, for example, a compositematerial comprising a PTFE/lead overlay with a bronze impregnatedinterlayer and a metal backing layer.

The bearing bushes may be directly mounted in the appropriate damperelement or may be mounted via a metal or plastics housing. Such housingsmay include bush retaining flanges and may be of greater axial extentthan the damper element in which they are mounted in order to allow theuse of longer bearing bushes which lowers bearing contact pressure. Thebearing bushes and/or housings may act as heat insulators to shield thepins to some extent against the worst effects of thermal expansion.

In an alternative arrangement the bushes may be axially retained byflanges provided on the element.

In certain applications needle bearings can be used instead of plainbearing bushes.

The pin is preferably hardened and is an interference fit in theassociated link which is made from softer material. Typically the linkwill be made from mild steel and the pin from hardened steel with groundend portions which engage in the bearing bushes.

In an alternative arrangement the pin can be made from stainless steeland the end portions can be turned (i.e. produced by a lathe turningoperation) to provide a helical patterned surface finish to assistbedding-in of the bearings.

If the pin is not made from harder material than the link the ends ofthe pin which engage the bearings are preferably of smaller diameterthan the central portion of the pin which is an interference fit withthe link in order to prevent damage to the pin end portions duringassembly of the pivot.

In certain applications, particularly if needle roller bearings areemployed, if may be desirable to hold grease captive within the bearing.

The invention also provides a torsional vibration damper comprising aninput element and an output element which are relatively rotatableagainst the action of a damping means which includes a plurality oflinkages, each linkage comprising one or more links mounted on oneelement via pivot pins and arranged to extend generally radially in useand connected with the other element by one or more additional links,the damper being characterised in that the generally radially extendinglinks are mounted on their pivot pins via gimbal pins which allow axialpivoting of the links relative to the pins.

BRIEF DESCRIPTION OF SEVERAL VIEWS OF THE DRAWINGS

The present invention will now be described, by way of example only,with reference to the accompanying drawings in which:

FIG. 1 is a view of a twin mass flywheel embodying the present inventiontaken in the direction B of FIG. 2;

FIG. 2 is a sectional view taken along the line 2—2 of FIG. 1;

FIG. 3 is a diagrammatic representation of a twin mass flywheel withassociated engine and gearbox;

FIG. 4 is a diagrammatic representation of an engine and associatedgearbox showing the engine crankshaft flexing;

FIG. 5 shows on a larger scale details of an alternative spaced bearingarrangement:

FIG. 6 shows details of a still further spaced bearing arrangement inwhich grease is held captive around the bearings by O-rings;

FIGS. 7 and 8 show details of a tolerance ring mounting arrangement fora bob-weight;

FIGS. 9 to 14 show details of yet further spaced bearing arrangements,and

FIG. 15 shows details of a gimbal-pin mounted bob-weight.

DETAILED DESCRIPTION OF THE INVENTION

With reference to FIGS. 1 ,2 and 3 there is illustrated a torsionalvibration damper in the form of a twin mass flywheel 10 which is formedby two damper elements 11 and 12. One flywheel mass 11 (also known asthe input flywheel mass) is fixed to a crankshaft 2 of an internalcombustion engine 1 by way of a central hub 20 and bolts 15. A frictionclutch 4 is secured to the second flywheel mass 12 (also known as theoutput flywheel mass) to connect the second flywheel mass with the inputshaft 5 of an associated gearbox 6.

The flywheel mass 11 comprises central hub 20, an input plate 21, acover plate 22, and a starter ring 23 which is welded to the input plate21. Cover plate 22 is secured at its outer periphery to input plate 21.Circlip 24 secures the inner race 51 of bearing 50 axially relative tothe hub 20. The input plate 21 is fixed to hub 20 by screws 25 prior toassembly onto the engine, and then also by the main bolts 15.

The second flywheel mass 12 comprises an output plate 30 a bearingretaining plate 31, and a pivot plate 32 all rotationally fast with eachother.

Under normal drive conditions and over-run conditions the twin massflywheel 10 as a whole rotates in a clockwise direction in the viewshown in FIG. 1 as indicated by arrow E. The engine crankshaft nominallyrotates about axis A and the gear box input shaft nominally rotatesabout axis B. Design of the engine, twin mass flywheel and gearboxassembly endeavours to ensure that axis A and B are co-linear. Howeverthis is not always the case, especially under running conditions.

Pivot plate 32 has an annular inner portion 32A with plurality of lugs32B which support pivots 43, a plurality of lugs 32C which inconjunction with abutments 33A form an end stop arrangement 33, and aplurality of lugs 32D, 32E which act on respective spring units 34D 34E.

Relative rotation between two damper elements 11 and 12 is controlled bya damping means which primarily comprises a plurality of pivotallinkages 40. The damping means also comprises the plurality of springunits 34D, 34E ,a friction damping device 60 and the plurality of endstop arrangements 33. All these components assist in controlling therelative rotation of the two damper elements 11 and 12 at specificrelative angular positions or in specific angular ranges.

Each pivotal linkage 40 comprises a generally radially extending link 41(also known as a bobweight link) pivotally mounted between a centre hubportion 35 of the output plate 30 and pivot plate 32 by way of a firstpivot 43, and an additional link 42 (in the form of a parallel pair ofarms 42A and 42B) pivotally mounted on the input flywheel mass 11 (byway of a second pivot 44). The links 41 and 42 are pivotally connectedto each other and bobweight 41 by means of a third pivot 45. It will benoted from FIG. 1 that the first pivot 43 is positioned radiallyinwardly of the second and third pivots 44 and 45.

The first pivot 43 is mounted in spaced bearings 43 _(a) and 43 _(b) inthe centre hub portion 35 and pivot plate 32 respectively.

Under no-load conditions with the clutch 4 disengaged, centrifugal forceacts on the pivotal linkages 40 and particularly on the first bobweightlink 41 and urges the linkages in a radially outward direction withpivot 45 adopting a position radially outboard of pivot 43 as shown inFIG. 1 (this position is regarded as the neutral position between thedrive and over-run directions of relative rotation of the damperelements). At higher rotational speeds the centrifugal force is greaterand whilst this does not affect the configuration under no-loadconditions it greatly affects the force required to move the flywheelmass 12 relative to the flywheel mass 11 i.e. the flywheel torsionalstiffness.

If the clutch is engaged and power is transmitted in the drive directionfrom flywheel mass 11 to flywheel mass 12 there is a tendency for thetwo masses to rotate relative to each other (flywheel mass 11 rotatesclockwise relative to flywheel mass 12 when viewing FIG. 1). Atrelatively low speeds when the influence of centrifugal force is smallerthe damper elements move readily relative to each other i.e. theflywheel torsional stiffness is relatively low. However at relativelyhigh speeds the influence of centrifugal force is much greater andrelative rotation of the damper elements requires greater force i.e. theflywheel torsional stiffness is relatively high. Thus the flywheeltorsional stiffness is speed sensitive.

If the clutch is engaged and power is transmitted in the over-rundirection from flywheel mass 12 to flywheel mass 11 the effects aresimilar to the above except that the direction of relative rotation isreversed (flywheel mass 11 rotates anticlockwise relative to flywheelmass 12 when viewing FIG. 1) and in the embodiment shown in FIG. 1 thefirst link 41 folds between the second link 42 i.e. between arms 42A and42B.

Input flywheel 11 is supported for rotation relative to output flywheel12 by bearing 50 which may be a self-aligning bearing, in this case adouble row self-aligning ball bearing which is held in position byretaining member 31 which is secured to output plate 30. Alternatively anon self-aligning bearing may be used.

During operation of the engine, the engine crankshaft 2 can flex, asshown diagramatically in FIG. 4. This flexing can be a first ordervibration (i.e. the crankshaft flexes once per revolution) second ordervibration (i.e. the crankshaft flexes twice per revolution, typicallycaused by the firing pulses in each cylinder in a 4 stroke engine) andalso higher and lower orders of vibrations are possible. Such flexingcauses the crankshaft flange 3 (and hence the attached input flywheel11) to be tilted out of plane A1 in which it normally generally lies bya relatively small but significant amount (X degrees) and into plane A2.Because of the complicated nature of the crankshaft flexing this tiltingmanifests itself as a tilting vibration which may take the form of aswashing and/or axial movement of the input flywheel.

However the transmission does not cause any such tilting vibrations orswashing in the output flywheel which continues to lie generally inplane B1.

Thus in conventional twin mass flywheels there is a ‘fight’ between theforces tilting the input flywheel 11 out of its plane A1 and the forceskeeping the output flywheel 12 in its true plane B1. This causesstresses in the twin mass flywheel components with the components ofeach flywheel mass tilting as they rotate by differing amounts dependingon the stiffness of the various components. Typically these stresses areseen in the components which link the two flywheels, that is the dampingarrangement and the bearing, thus reducing their service life.

However use of the self-aligning bearing 50 allows the input flywheel tolie in any plane as dictated by the flexing of the crankshaft, whilstalso allowing the output flywheel to continue to lie in plane B1. Thusany misalignment or tilting of the planes of the input and outputflywheels (A1, A2, B1) is accommodated in the self-aligning bearingrather than fought against. This reduces the stresses in the dampingcomponents and the bearing, thus increasing the service life of the twinmass flywheel. The range of tilting between the input and outputflywheels which the self-aligning bearing 50 accommodates is typicallyfrom 0.2 degrees to 3.0 degrees and is more typically 0.5 degrees.

In particular, once the self-aligning bearing allows the output flywheelto rotate true in the plane B1, the bob weight 41 which is mounted inthe output flywheel also runs true and is no longer forced to move backand forth axially as a result of tilting of the input flywheel. Thissignificantly reduces the stresses on the pivot between the bobweightand the output flywheel thus increasing its service life.

Belleville springs 50 a act to bias the planes of the input and outputflywheels (A1, B1) parallel to each other and this can be advantageousin some circumstances e.g. during balancing or assembly of the twin massflywheel.

Each generally radially extending bob weight link 41 has an intermediateportion 41 a of reduced axial thickness ‘x’ which allows the link toflex in an axial sense by an angle ∝ (typically in the range 0.2° to 5°)relative to the remainder of the flywheel as shown in dotted detail 41′in FIG. 2 as a result of the forces applied to link 41 due to tilting ofthe input flywheel 11 as shown in FIG. 4. If desired, the bob weightlinks 41 may not include the reduced thickness intermediate portion 41a.

Although the flywheel 10 described above is provided both with a mainself-aligning bearing 50 and the flexible links 41 the self-aligningbearing 50 may be omitted and the flexible links 41 may be used on theirown in certain applications or omitted.

In accordance with the present invention the first pivot pins 43 aresupported in spaced plain bearing bushes 43 a and 43 b. The pins 43 arepreferably a clearance fit in the spaced plain bearings bushes 43 a and43 b to ensure that any thermal expansion of pins 43 during use of theflywheel will not result in pins 43 binding in bearing 43 a and 43 b.Typically when a pin 43 of 10 mm diameter is used a clearance of 0.1 mm(100 microns) between the internal diameter of the bearings and theexternal diameter of the mounting pin is sufficient.

In the example shown, the bearings 43 a and 43 b are in the form ofpolymeric bushes (for example polyimide resin with a filler such asgraphite). Many other alternative bearing bush materials could be used,for example, a composite material comprising a PTTE/lead overlay with abronze impregnated interlayer and a metal backing layer.

Typically output plate 30, which is heated by clutch 4, is made fromcast iron and can be arranged to expand sufficiently during use of theflywheel to ensure that bearing bushes 43 a, 43 b do not distort tocause binding of the pins 43 in the bushes.

The bearing bushes are directly mounted in the hub portion 35 and pivotplate 32 but could alternatively, as shown in FIG. 5, be mounted via ametal or plastics housing 60 in either hub 35 and/or plate 32 suchhousings may include bush retaining flanges 61 and 62 and may be ofgreater axial extent than the flywheel component in which they aremounted in order to allow the use of longer bearing bushes which lowersbearing contact pressure.

In an alternative arrangement (not shown) the bushes may be axiallyretained by flanges provided on the flywheel.

In certain applications needle bearings can be used instead of plainbearing bushes.

The pins 43 are preferably hardened and are an interference fit in theassociated links 41 which are made from softer material. Typically thelinks will be made from mild steel and the pins from hardened steel withground end portions which engage in the bearing bushes 43 a and 43 b.

In an alternative arrangement the pins 43 can be made from stainlesssteel and the end portions can be turned on a lathe to provide helicalsurface patterning on the end portions to assist bedding-in of thebearings.

If the pin is not made from harder material than the link the ends ofthe pin which engage the bearings 43 a and 43 b are preferably ofsmaller diameter than the central portion of the pin which is aninterference fit with the link 41 in order to prevent damage to the pinend portions during assembly of the pivot.

In certain applications, particularly if needle roller bearings areemployed, it may be desirable to hold grease 68 captive within thebearing. This may be accomplished, as shown in FIG. 6, by providing anend cap 65 on the pin 43 and seals, such as O-rings 66 and 67, operativebetween the fink 41 and the adjacent hub portion 35 and pivot plate 32in grooves 66 a and 67 a respectively.

If desired the pins 43 may be provided with a central axial bore shownin dotted detail 70 in FIGS. 5 and 6. This bore minimises the effects ofthermal expansion on the pin by providing an inner periphery definingthe bore at which thermal expansion can take place and, when grease isused, allows the passage of grease from one end of the pin to the other.

The pins 43 may also be provided with coned end recesses 80 for theguidance of assembly probes during automatic assembly of the flywheelthus facilitating the lowering of the flange plate 32 into position onoutput plate 30.

FIGS. 7 and 8 show yet a further measure for accommodating movementbetween linkage pivots 43 and 44. In this arrangement the bob-weight 141is mounted on pivot pin 143 via a tolerance ring 141 b. which allowsbob-weight 141 to tilt relative to pin 143 through an angle β (typically0.2 to 5.0 degrees). Tolerance ring 141 b grips pin 143 so that pin 143pivots with bob-weight 141 within spaced bearing 143 a and 143 b.Tolerance ring 141 b also engages in a groove 143 c in pin 143 to locatepin 143 axially.

Instead of locating bearing 143 by locally deforming the plate 32 and 32a the bearing may be provided with an integral flange (similar to flange62 described above) which contacts plate 32.

FIG. 9 shows a bob-weight pivot arrangement in which the bob-weight 241is mounted on a pin 243 via bushes 243 a and 243 b. Bush 243 a ispressed into hub 35 and bush 243 b is pressed inside a cap or housing246 which is itself pressed into plate 32. Bushes 243 a and 243 b arelubricate by grease which under centrifugal action occupies the locatingshaded 247 in FIG. 9. A central opening 248 in cap 246 allows thepassage of an alignment tool during automatic assembly of the device. Nogrease exits via opening 248 during use of the flywheel.

A FIG. 10 shows a bob-weight pivot arrangement in which a bob-weight 341is mounted on a pin 343 via bearing bushes 343 a and 343 b of polymericmaterial. Bush 343 b is provided with an end cap portion 346 having acentral opening 348 for the passage of an alignment tool (shown indotted detail 349) during automatic assembly of the device. Alignmenttool 349 is offered up to hub 35 with plate 32, weights 341 and pins 343etc. to align pins 343 with bushes 343 a during assembly.

The bushes 343 a and 343 b are lubricated by grease which occupies thelocation shaded 347 when the flywheel is in use. The grease whichlubricates bush 343 b is injected after the removal of alignment tool349.

FIG. 11 shows a bob-weight pivot arrangement in which a bob-weight 441is mounted on a pin 443 via bushes 443 a and 443 b. Bush 443 a ispressed into hub 35 and bush 443 b is pressed into a sheet metal cap orhousing 446 which is itself pressed into plate 32. Bushes 443 a and 443b are lubricated by grease which under centrifugal action occupies thelocation shaded 447 in FIG. 11.

FIG. 12 shows a bob-weight pivot arrangement in which a bob-weight 541is mounted on a pin 543 via bushes 543 a and 543 b. Bush 543 a ispressed into hub 35 and bush 543 b is pressed into a sheet metal cap orhousing 546 which is itself pressed into plate 32. A plastics tubularcap 549 is push fit in a bore 550 in pin 543 and in a central aperture548 in cap 546. Cap 549 has axial grooves 551 which allow grease tomigrate to the locations shaded 547 in FIG. 12 under centrifugal actionwhen the flywheel is in use.

FIG. 13 shows a bob-weight pivot arrangement in which a bob-weight 641is mounted on a pin 643 via bushes 643 a and 643 b. Bush 643 a ispressed into hub 35 and bush 643 b is pressed into sheet metal cap orhousing 646 which is itself pressed into plate 32. Pin 643 has a centralzone 643 c which is part-spherical or barrel-shaped to allow weight 641to pivot axially relative to pin 643 as indicated by arrow P.Lubricating grease again occupies the shaded location 647 of FIG. 13when the flywheel is in use.

FIGS. 14 and 15 show a bob-weight pivot arrangement in which bob-weight741 is mounted on a pivot pin 743 by a gimbal pin 742 which allowsweight 741 to axially pivot relative to pin 743 as indicated by arrow Q.Pin 743 is supported in bushes 743 a and 743 b which may or may not belubricated by grease.

By mounting the pivot pins 43, 143, 243, 343, 443, 543, 643, 743 inspaced bearing bushes 43 a, 43 b: 143 a, 143 b: etc. the bearing contactpressure is lower as compared with an arrangement in which a singlebearing is used with the bob weight 41, 141, 241, etc. Also, a morestable support of the associated bob weight 41, 141, 241, etc. isprovided to resist the tendency of the bob weight to pivot axially dueto tilting of the flywheel masses as illustrated in FIG. 4.

A further benefit of the use of spaced bearing bushes 43 a, 43 b, etc.is that if these bushes are greased the grease tends to migrate to theradially outer zones of contact between the bearing bushes 43 a, 43 betc. and the associated pins 43, 143, etc. (as illustrated in, forexample, FIG. 9) where the migrated grease is indicated by the shadedlocation 247. It is these radially outer zones of contact where thehighest contact pressures occur due to the centrifugal forces generatedby the bob weights 41, 141, 241, etc.

As indicated above, bushes 43 a, 43 b and all the other bushes 143 a,143; 243 a, 243 b: etc. may be in the form of polymeric bushes (forexample polyimide resin with a filler such as graphite).

Many other alternative bearing bush materials could be used, forexample, a composite material comprising a PTFE/lead overlay with abronze impregnated interlayer and a metal backing layer.

If desired needle bearings, or similar rolling element bearingarrangement, can be used instead of any of the plain bearing bushes 43a, 143 b; 243 a, 243 b; etc.

As previously indicated, the present invention is not only applicable toa twin mass flywheel but is equally applicable to a torsional vibrationdamper where the input and output elements of the damper are ofrelatively light mass and are connected in a drive-line either on itsown or in combination with a fluid coupling as discussed above.

What is claimed is:
 1. A torsional vibration damper comprising an inputelement from which power is transmitted to an output element for onwardtransmission to a further drive line component via a damping means whichincludes a plurality of multi-links linkages, the input and outputelements are relatively rotatable against the action of the dampingmeans to absorb torsional vibrations and each linkage comprises one ormore links mounted on one damper element via pivot pins and arranged toextend generally radially in use and connected with the other damperelement by one or more additional links, the torsional vibrations damperbeing characterised in that the pivot pins of the generally radiallyextending links are supported from said one damper element via axiallyspaced bearings in said one damper element.
 2. A damper according toclaim 1 characterised in that the pins which mount the generallyradially extending links are a clearance fit in the bearings, the levelof clearance being sufficient to ensure that any thermal expansion ofthe pins during use of the damper is insufficient to cause the pins toblind in the bearings.
 3. A damper according to claim 2 characterised inthat the level of clearance is of the order of 1% of the diameter of thepins.
 4. A damper according to any one of claim 1 characterised in thatthe spaced bearings are in the form of plain bearing bushes.
 5. A damperaccording to claim 4 characterised in that the plain bearing bushes areof polymeric material.
 6. A damper according to claim 4 characterised inthat the bushes are in the form of a polymeric resin with a filler.
 7. Adamper according to claim 4 characterised in that the bushes are in theform of a composite material comprising a PTFE/lead overlay with abronze impregnated interlayer and a metal backing layer.
 8. A damperaccording to claim 7 characterised in that grease is held captive inareas of the contact between the pins and the bearings.
 9. A damperaccording to claim 8 characterised in that at least one of the spacedbearings or, where used, the bearing mounting housing is provided withan end cap portion which retains the grease.
 10. A damper according toclaim 9 characterised in that the end cap portion includes an aperturefor the passage of an assembly alignment tool.
 11. A damper according toclaim 10 characterised in that a cap with an axial passage therethroughfor grease is positioned in the aperture in the end cap portion.
 12. Adamper according to claim 10 characterised in that the pins which mountthe generally radially extending links each include bores or recessesfor the receipt of the assembly alignment tool.
 13. A damper accordingto claim 8 characterised in that the grease is retained by a sealoperative between the generally radially extending link and said onedamper element.
 14. A damper according to claim 1 characterised in thatat least one of the spaced bearings is mounted in said one damperelement via a metal or plastics housing.
 15. A damper according to claim14 characterised in that the housing includes a first flange to retainthe bearings within the housing.
 16. A damper according to claim 15characterised in that the housing includes a second flange to retain thehousing within said one damper element.
 17. A damper according to claim14 characterised in that the housing has a greater axial extent thansaid one damper element in which it is mounted.
 18. A damper accordingto any one of claim 1 characterised in that the axially spaced bearingsare designed to act as heat insulators to shield the pivot pins from theeffects of thermal expansion.
 19. A damper according to claim 1characterised in that the pins are in an interference fit in thegenerally radially extending links, the links being made from a softermaterial than that of the pins.
 20. A damper according to claim 1characterised in that the pins which mount the generally radiallyextending links each include an axially extending through bore.
 21. Adamper according to claim 1 characterised in that of the pins whichmount the generally radially extending links have part-spherical orbarrel-shaped zones to allow axial pivoting of the generally radiallyextending links relative to the pins.
 22. A damper according to claim 1characterised in that the generally radially extending links are mountedon their pivot pins via tolerance rings which allow axial pivoting ofthe generally radially extending links relative to the pins.
 23. Adamper according to claim 1 characterised in that the generally radiallyextending links are mounted on their pivot pins by gimbal pins whichallow axial pivoting of the generally radially extending links relativeto the pins.
 24. A damper according to claim 1 characterised in that thegenerally radially extending links have a concentration of mass at theirradially outer ends thus operating as bobweights which damp the relativerotation of the input and output damper elements.
 25. A damper accordingto claim 1 characterised in that the generally radially extending linksare designed to flex in an axial sense to accommodate forces applied tothe generally radially extending links as a result of relative tiltingof the input and output damper elements.
 26. A damper according to claim1 characterised by being in the form of a twin mass flywheel where theinput and output elements comprise input and output flywheel massesrespectively.
 27. A damper according to claim 1 characterised by beingconnected in combination with a fluid coupling.